The present invention generally relates to bearings and more particularly to bearings of the type that are dynamically lubricated, wherein the bearings are configured to inhibit viscous heating of the lubricant and thereby operate at relatively lower temperatures.
FIG. 1 schematically represents a high-bypass turbofan engine 10 of a type known in the art. The engine 10 is schematically represented as including a nacelle 12 and a core engine module 14. A fan assembly 16 located in front of the core module 14 includes an array of fan blades 18. The core module 14 is represented as including a high-pressure compressor 22, a combustor 24, a high-pressure turbine 26 and a low-pressure turbine 28. Air is drawn into the inlet duct 20 of the engine 10 and then compressed by the compressor 22 before being delivered to the combustor 24, where the compressed air is mixed with fuel and ignited to produce hot combustion gases that pass through the turbines 26 and 28 before being exhausted through a primary exhaust nozzle 30. To generate additional engine thrust, a large portion of the air that enters the fan assembly 16 is bypassed through an annular-shaped bypass duct 32 before exiting through a fan exit nozzle 34.
FIG. 1 schematically represents the high-pressure compressor 22 and high-pressure turbine 26 as mounted on the same shaft 36 so that the flow of hot exhaust gases that pass through the high-pressure turbine 26 turns the turbine 26 as well as the compressor 22 via the shaft 36. The shaft 36 is supported with multiple rolling element bearings, of which a ball bearing 38 is represented in FIG. 1 located near the entrance of the compressor 22. The shaft 36 is mounted within an inner race of the bearing 38, while an outer race of the bearing 38 is supported by a static structure of the core engine module 14. From FIG. 1, it should be apparent that the axis of the bearing 38 coincides with the centerline 35 of the engine 10.
FIG. 2 represents a cross-sectional view of a portion of the bearing 38 of FIG. 1. As a ball bearing, the bearing 38 is shown as comprising an inner race 40, an outer race 42, rolling elements (balls) 44 (of which only one is shown in FIG. 2), and a cage 46. The rolling elements 44 reside within grooves 50 and 52 defined in opposing surfaces of the races 40 and 42, respectively, such that in combination the grooves 50 and 52 define the load-bearing contact surfaces of the bearing 38. The cage 46 serves to maintain separation between the rolling elements 44. In FIG. 2, each groove 50 and 52 is represented as having a semi-spherical cross-sectional shape that closely matches the curvature of the rolling elements 44, though with a slightly larger radius than the rolling element 44. Such a shape is commonly referred to as a circular arch, and provides a single contact point between each rolling element and each individual race 40 and 42. The contact points or patches 54 and 56 are diametrically opposed as schematically represented in FIG. 2. The term “patches” refers to the fact that a true point contact does not exist when a bearing is loaded, and that the contact patches 54 and 56 have elliptical shapes caused by loading between the rolling elements 44 and the races 40 and 42.
Due to the high rotational speeds required of the shaft 36, the bearing 38 must operate at high rotational speeds. Specifically, though the outer race 42 does not rotate, the inner race 40 rotates at the same speed as the shaft 36 and the rolling elements 44 therebetween rotate around the inner race 42 at a lower speed than the inner race 42. High-speed ball bearings of the type represented in FIG. 1 are often dynamically cooled with a lubricant that flows through the bearing 38. In FIG. 2, the inner race 40 of the bearing 38 is provided with under-race lubrication features in the form of multiple inlets 48 through which a lubricant (typically oil) is introduced into an annular-shaped cavity 58 defined by and between the inner and outer races 40 and 42 of the bearing 38. The lubricant provides both lubrication and cooling of the rolling elements 44 and cage 46 within the cavity 58. Under the influence of centrifugal forces caused by the spinning inner race 40, the lubricant supplied through the inlets 48 flows radially outward to contact the cage 46, the rolling elements 44, and the outer race 42. As represented in FIG. 2, because the bearing 38 is provided with an under-race lubrication system, the cage 46 is typically configured so that it bears against cage lands 60 on the inner race 40.
Because the outer race 42 does not rotate and the inner race 40, rolling elements 44 and cage 46 are moving at different speeds, the lubricant within the cavity 58 tends to churn, which as used herein refers to nonhomogeneous flow patterns within the cavity 58. Analysis has shown that churning primarily occurs at the outer race 42, and more particularly within the groove 52 of the outer race 42, where the lubricant tends to accumulate before exiting the bearing 38. Analysis has also indicated that churning occurs between the cage 46 and inner race 40, as a result of a low pressure area created by the rotational effects of the high-speed rolling elements 44. In conventional dynamically-lubricated bearing designs, the lubricant exits the bearing 38 at the inner and outer diameters of the cage 46 on both axial ends 62 and 64 of the bearing 38, with the majority of the lubricant exiting at the outer diameter of the cage 46 in view of the position of the cage 46 against the inner race cage lands 60. Furthermore, when the bearing 38 is operating with an axial load (as represented in FIG. 2), a majority of the lubricant will exit at the outer diameter of the cage 46 and on the unloaded side of the bearing 38.
Various approaches have been proposed to promote the lubrication of rolling element bearings, including efforts to reduce heat generation at high rotational speeds. One such approach disclosed in U.S. Pat. No. 5,749,660 to Dusserre-Telmon et al. is the inclusion of a drain feature in the outer race. The drain features are orifices having entrances that are located in the groove of the outer race and exits that are located on the outer circumference of the outer race, so that the lubricant drains from the bearing by flowing completely through the outer race in a radially outward direction. The grooves of the inner and outer races do not have circular cross-sectional shapes matching the curvature of the rolling elements, but instead are described as having rather conical shapes that define vertices which form part of a central circumference of each groove. As a result, the rolling elements never cover the drain orifices located in the outer race groove, but instead touch the outer race at two lateral contact patches on each side of the orifices. Such a configuration is similar to conventional bearing races that have what is commonly referred to as a gothic arch, in which case the race is defined by two radii with different axes of curvature, as opposed to the aforementioned circular arch defined by a single radius. Similarly, the rolling elements contact the inner race groove at two lateral contact patches on each side of inlet orifices that are formed in the inner race to introduce the lubricant into the bearing, with the result that each rolling element can have as few as two and as many as four contact points with the inner and outer races.
While not intending to promote any particular interpretation of U.S. Pat. No. 5,749,660, it appears that the four-point contact may not be capable of operating with a low axial load conditions that would occur when the rotor thrust load changes direction during transitions from low to high speed conditions, as would be required in most gas turbine applications of the type represented in FIG. 1. Furthermore, the drain orifices may contribute significant stress concentrations in the outer race and reduce the ability of the bearing to survive ultra-high load events, such as fan blade out conditions.